Analytical design and computational modelling of cooling system
for combat vehicle using porous medium technique
Abstract
The present project is aiming at analytically carrying out numerical calculations required for
the designing of a compact heat exchanger for a combat vehicle’s power pack. Since heat
signature and volumetric space constraints are the biggest challenges for a combat vehicle thus
it is extremely important to reduce the vehicle thermal signature in order to maintain its stealth
behaviour. Moreover, a combat vehicle has to be always ready and it should be able to work
even under degraded conditions, this whole project is based on keenly picking out
characteristics required for a combat vehicle pertaining to various geographical features.
Furthermore, the designed cooling pack is to be correlated with the help of computational fluid
dynamics simulations using porous media technique.
Introduction
Heat exchanger is one of the most important equipment, it plays a major role in the modern
day industry be it petrochemical, automobile, aerospace, space industry, electronics or power
industry etc. The use of heat exchangers in the automotive industry dates back to 1910s, earlier
designs were simpler due to limited manufacturing technology. However, now owing to the
phenomenal advancements in the design, analysis and manufacturing technology, automotive
heat exchangers have become more efficient and compact. An automotive heat exchanger or
simply radiator is an integral part of the vehicle and is the backbone of the engine compartment.
It not only maintains the temperature of the engine in the operable range, but it provides other
functionalities too viz. heating the passenger compartment in cold climate for comfortable ride,
cooling the transmission fluid using oil coolers and with the development of turbocharging and
supercharging to obtain additional power from the same engine by increasing the density of the
inlet air, means pressurizing inlet air that will increase the air temperature hence came the
charged air coolers in existence. Numerous advancements have been made in the heat
exchanger design methods, the most quintessential requirement today is of reducing the size,
cost and increasing the reliability.
For improving the heat transfer characteristics of the heat exchanger it is a common practice
to increase the surface area using fins and to increase the turbulence using turbulators. In the
present day, the use of compact heat exchangers are gaining momentum because they provide
a good blend between size and efficiency as compared to the traditional counterparts.
Military combat vehicles such as tanks are designed for combat situations and special
missions. They are provided with heavy armored protection, thus the engine and the other sub-
assemblies are completely enclosed inside the vehicle’s hull and are ventilated through ballistic
grilles making it challenging for the air to flow. The space inside a combat vehicle is also
limited and the heat rejection rate is enormously high. This makes design of cooling system for
a combat vehicle challenging.
With the use of advanced computational fluid dynamics, designing the heat exchanger for
combat vehicle will prove promising and cooling system with higher cooling capacity can be
made. The basic techniques available to analyze the heat exchangers are log mean temperature
difference (LMTD) and effectiveness-NTU (ε-NTU) method. Barbaros et. al [1] states solely
relying on implementation of LMTD and ε-NTU is not feasible especially for vehicle radiators
which may include custom designed fin configuration. CFD analysis of radiator is not feasible
due to the requirement of extremely high number of cells to solve the complex nature of
geometry. Thus, the fluid flow through fins can be modelled using a porous media, this will
lower the computational complexities and the porous media model does not require boundary
layer meshing since the friction and heat transfer parameters are already included in the porous
parameters. [2]
Compact heat exchangers
The process of heat exchange between two fluids that are at different temperatures and
separated by a solid wall occurs in many engineering applications. The device used to
implement this exchange is termed as a heat exchanger. Decreasing size and increasing heat
load is the typical feature of the modern-day heat exchanger industry. A compact heat
exchanger is generally defined as one which incorporates a heat transfer surface having a high
“area density”. Quantitatively, a compact heat exchanger is one that has an area density (heat
transfer surface area to volume) greater than 700m
2
/m
3
as represented in Figure 1 below.
Typical compact heat exchanger applications are automotive, aerospace and cryogenics, with
continuing demand of higher heat transfer with further shrinking of available space, many new
areas coming up in the use of compact heat exchangers including cooling of electronic
equipment, cooling of laser and related technologies and cooling technology for fuel cells.
Porous modelling
The porous media model incorporates an empirically determined flow resistance in a region
of your model defined as "porous''. It is governed by three models. The simplest model is the
Darcy’s model which is given for laminar flow through the media:
(1)
where, Δp is the pressure drop, l is the pipe length, V is the average velocity, μ is the dynamic
viscosity and α is permeability of porous medium. The other two models Forchheimer and
Forchheimer-Brinkman model are applicable in case of turbulent flows as pressure drop and
velocity’s relation becomes non-linear.
In case of porous modelling, the pressure drop across the ends is given using a polynomial
of velocity.
(2)
The coefficients C
1
,C
2
and C
3
are calculated using curve fitting the heat exchanger’s pressure
drop vs volumetric flow rate graph and further the force terms are calculated based on the
pressure drop data and further the source terms are computed by dividing the force with volume
Figure 1- Heat transfer surface area density spectrum of exchanger surfaces [3]
of the porous media. Source terms are required because the CFD code implements this source
term locally in the generated cells. [4]
Aims and Objectives
This project is aimed at analytically designing and simulating the cooling system for combat
vehicle using porous medium technique.
In order to achieve the aforesaid aim the objectives are as follows-
1. To conduct extensive study on various combat vehicle’s cooling pack requirements,
configurations and technology adopted.
2. To develop a tool for sizing of radiator cores, mass flow rate and pressure drop
calculations.
3. To achieve a compact heat exchanger having the ability to cool the combat vehicle
effectively along with using comparatively less amounts of space.
4. To design 3D CAD model of cooling pack.
5. To perform a flow simulation analysis using porous medium for ensuring appropriate
mass flow rate.
6. To optimize the shroud shape for minimal flow resistance and required flow rate.
7. To compare the use of multiple electric fans vs hydraulic fans.
Literature review
Project background
We have reached at that stage of technological advancement in automotive domain yet we
have not been able to master the heat generation and dissipation from engines completely, we
currently have about 35 – 40 % efficient engines available rest of the energy generated gets
exhausted. The world is now currently in the need of sources of energy as renewable sources
are limited and might exhausted soon in the upcoming near future, thus harnessing ever bit of
energy generated is the major concern in every domain nowadays. We now have the advantage
of high computational power that can be used as a head start in designing optimum and lean
systems.
Military vehicles generate comparatively greater amounts of heat than commercial vehicles.
Shah’s [5] work suggest that a military vehicle has at least the following components extra that
count for greater heat signature viz. oil coolers, evaporators, condensers and charged air
coolers. Also they have to work in combat zones with tense environments, hence the need of a
high performance cooling pack is essential. In case of military vehicles space is the biggest
constraint, the design should have the best performance without taking up much space.
Previous work
Use of CFD in designing fluid systems is not a new concept. It has been used since its
inception in the early days. A lot of industries have successfully been able to develop accurate
CFD models. The major issue in developing a correct CFD model is that it is extremely time
consuming and costly as the surfaces becomes compact, the time and probability of solver to
converge reduces. Thus, porous modelling is preferred in such cases. Carluccio et.al [6] have
worked on similar grounds on heavy earth movement vehicles by schematizing complex fin
geometries as porous medium. Mao and Cheng [7] have used porous modelling in their work
and they conclude that porous medium method can effectively simulate fluid flow and heat
transfer in the radiator of a heavy-duty cooling system, and provide a reliable and economic
tool for industrial applications. Cetin et.al [8] suggests that analysis of a radiator by using
porous medium approach gives reasonable and reliable results. By using CFD analysis, design
cost may be decreased dramatically by easing the experimental testing process. Porous
parameters of a given radiator plays a significant role in distribution of flow field across the
exchanger and required as a boundary condition for CFD.
These parameters of porous medium can be obtained using classical/empirical equation or
via conducting a pressure based simulation of fins using CFD.
Comparative work-
The table presented below is of 8 x 8 wheeled armoured carriers in the world, by listing them
along with their engine rating, we get an initial assessment of the cooling load in such class of
military vehicles. These vehicles are designed for an ambient temperature ranging from -25 to
+50
o
C, and are powered by heavy duty engines.
Table 1- Benchmarking of wheeled armoured vehicles
S.No.
Vehicle
Country of
Origin
Engine
Engine
rating
1
Patria AMV
Finland
DI 12 Scania
600 HP
2
Mowag Piranha V
Switzerland
MTU 6V199 TE21
550 HP
3
Pandur II
Austria
Cummins ISLe HPCR
450 HP
4
Kodiak
Canada
Caterpillar 3126
350 HP
5
Dragoon 300
USA
Detroit Diesel 6V-53T
300 HP
6
Boxer
Germany
MTU 8V199 TE20
711 HP
7
VPK-7829 Bumerang
Russia
BarnaulTransMash UTD-32TR
510 HP
8
WhAP
India
Cummins ISXe
600 HP
A combat vehicle has to protect itself from heat seeking systems which makes it difficult to
place the cooling system in the front of the vehicle. Furthermore, they are provided with
amphibious fording characteristics which allows them to tackle marches and swamps. Hence,
the radiator of military combat vehicles are generally placed in the top of their hull, such a
placement ensure fording capability and reduces ground envelope of heat signature. [9]
(a)
(b)
Figure 2-Radiator placement in military combat vehicle (a) Patria AMV, M1117
armoured vehicle [10]
Analytical design
This section gives detailed information about compact heat exchanger design, it deals with
rating and sizing of the heat exchanger. The cooling pack rating is concerned with the
determination of heat transfer rate, fluid outlet temperatures for prescribed fluid flow rates, inlet
temperatures and allowable pressure drop for the heat exchanger. On the other hand, the sizing
of radiator, involves dimensioning of heat exchanger by selecting an appropriate inlet/outlet
temperatures, flow rates, and pressure drops. In the present work, effectiveness-NTU method
has been used as suggested by Kays & London [11].
Figure 3- An example of radiator location in military vehicle [9]
Introduction
The cooling system for a combat vehicle is one of the most critical sub-systems of all as it
has direct impact on various sub-system and assemblies of the vehicle. Combat vehicles are
designed to work in the close-proximity with enemy, therefore the components are enclosed
inside a hull structure made up of ballistic grade composite materials. Moreover, unlike
commercial vehicles a combat vehicle’s cooling pack has other additional components to cool
down its fire control system, Navigation system and sighting system etc. Basically a combat
vehicle cooling pack has oil coolers for engine, transmission, hydraulic actuators, steering,
winches, charged air coolers and HVAC condenser.
During combustion of Heavy duty Diesel engine, the peak temperature of over 2000 °C is
reached, so overheating of components must be prevented by an appropriate cooling
system.[12] The demand for cooling pack is combat vehicle is essential :-[13]
To promote a high volumetric efficiency
To ensure proper combustion
To secure mechanical operation and reliability
A combat vehicle is provided with louvers at air intakes to protect the vehicle form missiles,
the presence of such louvers poses hindrances the air path which makes it demanding to cool
down the engine and associated sub-assemblies. These factors make the cooling system design
process of combat vehicles quite different as compared to commercial vehicles.
Theory
The theory of a cooling system aims at analysing the effect of fluid flow, evaluation of
convective heat transfer, system resistances, flow characteristics and selection of fins.
Mechanism of cooling
Automotive radiators consist of two header tanks (inlet and outlet), which holds the cooling
fluid, and a core. The core of a radiator comprised of tubes/plates and fins. The tubes/plates run
lengthwise from header to header and the fins are located in between the tubes. The fins serve
the purpose of increasing the heat transfer area of the radiator without changing the mass flow
rate of air across the radiator. When the radiator is operating, coolant flows through the tubes
and airflows through the core of the radiator. This airflow through the radiator lowers the
temperature of the coolant via the principle of cross-flow convection.
Figure 4- (a) Plate fin compact heat exchanger, (b) Tube fin compact heat exchanger. [14]
Heat transfer analysis
A basic heat transfer analysis for heat exchanger given by Shah [5] is presented here-
The heat exchanger effectiveness, ε is defined as-
(3)
where, C
min
is the smaller of C
c
and C
h
and the product mass flow rate and specific heat
capacity (mC
p
) is often denoted by C.
By definition, the effectiveness, which is dimensionless, must be in the range 0 ≤ ε ≤ 1. This
is useful because if ε,
and
are known, the actual heat transfer rate may readily be
determined from the expression-
(4)
For any heat exchanger, it can be shown that effectiveness is a function of capacity ratio and
number of heat transfer units (NTU). (Kays & London [11])
Number of transfer units (NTU) is a dimensionless parameter that is widely used for heat
exchanger analysis and is defined as-
(5)
Where, U and A are overall heat transfer coefficient and overall heat transfer area
respectively, for the radiator that would be used.
The overall heat transfer coefficient is defined as-
(6)
Pressure drop and System resistances
When air passes through the radiator, it creates pressure drop across the core due to fin
friction and air turbulence. The interaction of fins with the air constructs and destructs multiple
boundary layer lead to shift in air path and creates turbulences. Flow properties are designed
based on factors including colburn, fanning friction and their relationships with heat transfer
coefficients, which were extensively studied and tested by Kays & London [11].
The hydraulic diameter is given by-
(7)
The mass velocity G is given by-
(8)
The Reynolds number is given by-
(9)
The convective heat transfer coefficient h in terms of the Colburn factor j is given as-
(10)
The basic performance data for an enhanced fin surface is often shown as curves of the
Colburn factor and the Fanning friction factor, plotted versus Reynolds number. Kays and
London [11] presents j and f vs R
e
for a large number of compact surfaces, in one of the first
comprehensive collections of data on enhanced surfaces for compact heat exchangers.
The core pressure drop can be calculated as [11]-
(11)
where,
ρ
i
,ρ
o
, ρ
m
are density at inlet & outlet and the average density evaluated at the average
temperature between the inlet and outlet respectively.
K
C
= entry term coefficient
K
e
= exit term coefficient
The first term inside the bracket shows the entrance effect, the second term shows the flow
acceleration effect, the third term shows the core friction, and the last term represents the exit
effect.
In Equation 11, the frictional pressure drop generally dominates and accounts for about 90%
or more of the total pressure drop across the core. The friction factor f, has been found
experimentally and plotted for various surfaces. The entrance and exit losses coefficient K
c
,
and K
e
, values are given by Kays & London. [11]
Heat transfer surfaces
A wide range of fin geometries have been used to obtain enhanced heat transfer, new and
more effective enhanced surfaces are also being developed. Typical fin spacing are 400 to 1000
fins per metre. Due to their small hydraulic diameter, these surfaces are usually operated in the
Reynolds number ranging between 500 –10,000. As a result, they are effective in the low
Reynolds number range. A high-performance surface should have the characteristics that will
enhance the heat transfer of a heat exchanger, without incurring penalties on friction and
pressure drop.
Figure 5- Heat exchanger core with different fin surface geometries (a) Offset-serrated fins,
(b) Dimpled fins, (c) Corrugated fins, (d) Louvered fins, (e) Perforated fins, (f) Pin fins [10]
Methodology
The thermal design methodology to estimate the heat transfer performance as well as core
sizing are given in the flowchart. Based on the following methodology a worksheet using MS
Excel has been developed to calculate various performance parameters of radiator. This
worksheet has been made with the intention to simulate different conditions and enable value
iterations when necessary in some calculations.
Figure 6- Heat exchanger design methodology [14]
Design parameters of cooling pack
The cooling pack was designed by selecting a 450 HP engine. Engine selection and cooling
load calculations are presented in section 3.5. The cooling pack comprises of two radiators:
Main and Auxiliary radiators. Main radiator is meant for cooling the dynamic components of
the vehicle e.g. Engine and Transmission etc. whereas the auxiliary radiator is designed for
auxiliary heat loads of the vehicle. The cooling load for the main radiator is 276 kW and for
auxiliary radiator is 142 kW. The fins taken in consideration are of aluminium grade 3000 series
(thermal conductivity-336 W/m.K). The coolant considered for calculation is 50-50% ethylene-
glycol.
The following parameters based on initial calculations are taken in the overall design process
of the cooling pack-
Table 2- Design parameters of the cooling pack
Auxiliary radiator
Air side (cold fluid)
Coolant side (hot side)
Inlet mass flow rate
9.8 kg/s
1.66kg/s
Inlet flow temperature
50
o
C
112.8
o
C
Main Radiator
Air side (cold fluid)
Coolant side (hot side)
Inlet mass flow rate
9.8 kg/s
5.34 kg/s
Inlet flow temperature
66
o
C
111.1
o
C
The thermophysical properties considered for designing are presented in the following
tables-
Table 3-Thermophysical properties of 50% ethylene glycol solution
Temperature
Specific Heat
Dynamic
Viscosity
Prandtl Number
Density
(
o
C)
C
p
(kJ/kgK)
( kg/m s)
(kg/m3)
0
3.18
0.01029
86.3
1083
20
3.31
0.00459
47.6
1072
40
3.42
0.00238
20.1
1061
60
3.52
0.00139
11.8
1048
80
3.59
0.00099
8.3
1034
100
3.65
0.0008
6.6
1020
120
3.68
0.00066
5.4
1003
Table 4- Thermophysical properties of air at atmospheric pressure
Temperature
Specific Heat
Dynamic
Viscosity
Prandtl Number
Density
(K)
C
p
(kJ/kgK)
(10
-5
kg/m s)
(kg/m3)
275
1.0038
1.725
0.713
1.284
300
1.0049
1.846
0.707
1.177
325
1.0063
1.962
0.701
1.086
350
1.0082
2.075
0.697
1.009
375
1.0106
2.181
0.692
0.9413
400
1.0135
2.286
0.688
0.8824
450
1.0206
2.485
0.684
0.7844
500
1.0295
2.67
0.68
0.706
550
1.0398
2.849
0.68
0.6418
600
1.0511
3.017
0.68
0.5883
650
1.0629
3.178
0.682
0.543
700
1.075
3.332
0.684
0.5043
750
1.087
3.482
0.687
0.4706
800
1.0987
3.624
0.69
0.4412
850
1.1101
3.763
0.693
0.4153
900
1.1209
3.897
0.696
0.3922
950
1.1313
4.026
0.699
0.3716
Engine selection and heat load calculations
From the study of various combat vehicles and their intended running environment the
engine with the following specifications was selected-
Engine cooling load calculations-
The cooling load is the amount of heat that needs to be dissipated by the heat exchanger.
The heat supplied to the engine is through the energy in fuel (calorific value or heating value)
and is given by-
(12)
where,
m
f
= mass flow rate of fuel
C
v
= calorific value of fuel
Table 5- Technical specifications of Caterpillar C11 ACERT engine
Engine specifications
Engine
Caterpillar C11
Dimensions
1202.06 x 1055.9 x 1176.0 mm
Configuration
In-line, 6-cylinder
Aspiration
Turbocharged \ After cooled
Displacement
11.1 L
Bore & Stroke
130 x 140 mm
Fuel System
Mechanical Electronic Unit Injector
Horsepower
450 HP (336kW) @ 2100 RPM
Torque
2056 Nm @ 1400 RPM
Emission
standard
EU STAGE IIIA with Advanced Emission Reduction Technology
(ACERT)
Weight
930 Kg
Figure 7- Engine data sheet of Caterpillar C11 ACERT
Table 6- Fuel rate vs engine speed
Engine Speed RPM
Fuel Rate L/hr
2100
87.8
2000
86.7
1900
86.5
1800
86.2
1700
84.9
1600
83.3
1500
79.8
1400
77.1
1300
69.1
1200
64.5
1100
55
Since, in an IC engine combustion process, the cooling of burned gases to room temperature
is not possible and therefore the water vapour is not condensed, hence lower heating calorific
value is to be considered.
A combat vehicle has to operate mostly in a region of maximum torque. Thus, cooling load is
calculated based on the above marked/dotted box conditions.
The thermal efficiency of engine is given by-
(13)
Maximum torque occurs at 1400 RPM, hence
~ 36%
Lower heating calorific value is given by-
(14)
Now, engine heat load is
The following table is based on the experimental and past data acquired in the same line.
Table 7- Heat load percentage contribution of various components of combat vehicle
Heat Load
% contribution
Wattage (kW)
Engine
18%
151
Engine Oil cooler
7%
58
Transmission Oil cooler
8%
67
Intercooler
8%
67
Fuel Cooler
1%
8
Hydraulic Oil cooler
5%
41
Steering oil cooler
3%
25
Total heat load for Main radiator
276 kW
Total heat load for Aux radiator
142 kW
Cooling system layout
The cooling system layout is presented in the figure (8). The thermal circuit comprises of
various coolers namely transmission oil cooler, engine oil cooler, multistage intercooler,
hydraulic oil cooler, steering oil cooler, and fuel cooler. The main and auxiliary radiators are
designed to dissipate the heat generated by the said sources in the vehicle. Most of the small
coolers designed are brazed plate heat exchangers, and main and auxiliary radiators are plate-
fin compact heat exchangers.
Figure 8- Schematic showing the cooling pack thermal circuit
Computational modelling
Introduction
• Computational modelling is the utilization of workstation to simulate and concentrate
on complex frameworks utilizing math, physical science and software engineering.
• Computer modelling allows engineers and scientists to perform numerical
experimentation of simulated physical condition with the aid of a computer.
• A computational model contains a variety of factors that describe the framework in
question. Simulation is finished by adjusting the variables alone and observing the
effects.
• Computer modelling allows scientists to conduct lots of simulated experiments with
the aid of computer.
Computational fluid dynamics is the analysis of systems involving fluid inflow and heat
transfer by means of computer- aided simulation. The approach is very powerful and spans a
huge range of business and non-industrial application areas. The availability of affordable
excessive performance computing hardware and the creation of consumer- pleasant interfaces
have led to the improvement of industrial CFD applications. Historically only Experimental
Fluid Dynamics (EFD) and Analytical Fluid Dynamics (AFD) were possible but with the
advent of digital computer and advancing with improvements of computer resources CFD is
made possible. CFD is carried out to an extensive variety of studies and engineering problems
in many fields which includes aerodynamics, aerospace, hypersonic, climate simulation,
environmental engineering, industrial device design and evaluation, organic engineering, fluid
flows and heat-transfer, engine combustion analysis, and visual effects for movie and video
games.
Today, well-tested commercial CFD packages have not only made CFD analysis a routine
design tool in industries, but have also helped the engineer’s in designing physical systems
more efficiently. All formal CFD software contains: (i) pre-processor (ii) main solver and (iii)
post processor
Discretization method
Discretization is the process of moving continuous functions, variables, models, and
equations into discrete counter parts in applied mathematics. This technique is usually achieved
as a primary step in the direction of making them suitable for numerical assessment and
implementation on virtual computers.
There are 3 distinct methods of numerical solution techniques: finite difference, finite
volume and finite element methods. A reason in each is to convert the differential equations
into algebraic equations. The primary differences between the 3 strategies are associated with
the way the differential equations are converted to algebraic equations.
Finite Difference Method
• Governing equations in differential form→ domain with grid→ replacing the partial
derivatives by approximations in terms of node values of the function
• It is often used, classic method
• Applied to structured grids
Finite Volume Method
• Governing equations in integral form→ solution domain is subdivided into a finite
number of contiguous control volumes→ conservation equation applied to each
control volume.
• It is famous because of conservation & structured and unstructured meshes .General
transport equation for mass, momentum, energy, etc. is applied to each cell and
discretize
• Applied to any type of grids, especially complex geometries.
Finite Element Method
• It is a numerical tool for determining approximate solutions to various problems
• This is a popular method for numerically solving differential equations arising in
engineering and mathematical modelling
• Typical problem areas of interest include the traditional fields of structural analysis,
heat transfer, fluid flow
Governing Equations
The governing equations of the fluid motion depends on three fundamental principles of
physics i.e mass, momentum, and energy conservation. These principles express that mass,
momentum, and energy are steady constants inside a closed system. The examination of fluid
stream with temperature variation depends on specific physical properties.
The equation for the conservation of mass is specified as-
(15)
where,
= density
= velocity
= gradient operator
Conservation of momentum which can be referred to as the Naiver-Stokes equation is-
(16)
where,
p=static pressure
= viscous stress tensor
= gravitational force per unit volume
The first term represents local change with time, second deals with momentum convection,
third one is surface force, fourth term depicts diffusion term and the final term is mass force.
Conservation of energy is given as-
(17)
where,
dQ = heat added to the system
dW = work done on the system
dE
t
= increment in the total energy of the system
CFD Procedure
Figure 9- Flowchart representing CFD simulation procedure
CAD modelling
The following steps have been followed for designing the CAD model-
1. Sketches were created based on the dimensions provided by the analytical
design
2. Appropriate features, such as extrudes and fillets were selected as per design
requirements
3. All the designed components were assembled.
Solidworks 2022 SP.1 has been used extensively for modelling of cooling pack components
such as Louvers for radiator housing, fan housing, radiator core, and cowl for directing air
outlet.
Few design considerations that must be followed while designing the cooling pack are-
1. The air inlet section before radiator has to be made diverging. The purpose is to
reduce velocity of air. This increases effective time of heat rejection. Reducing
velocity also reduces drag force.
2. The fan inlet and the radiator outlet must have enough space between them to
reduce the chances of wake formation and to obtain the entrance length
promoting fully developed flow inside the shroud.
3. The fan outlet ought to have a converging section to promote increase in
velocity of the hot exhaust air such that it matches the free stream velocity. This
reduces turbulence and maintains optimal pressure drop/mass flow rate.
4. Sharp curves have to be avoided to ensure less drag and recirculation.
5. Proper partition/ baffle plates should be provided between fan outlet louvers and
inlet air louvers to avoid flow recirculation back into the inlet.
6. The air inlet area should be 80% of the frontal area of the radiator.
Figure 10- Image showing developed CAD model (a) Side-view of the radiator assembly, (b) Top-view of the assembly showing louvered
grilles, (c) Front-view of the assembly showing curvature of the cowl, (d) Isometric view of the model
Numerical simulation
The simulation was carried out using Dassault Systèmes Solidworks Flow Simulation
Package using Finite volume method for solving the Naiver Stokes equations. Two models
were made one with hydraulic fans and another with electric fans. The number of cells
generated were 416636 in the mesh. Solidworks Flow Simulation solves Reynold Average
Naiver Stokes equations (RANS) with K-Epsilon turbulence model and standard wall
functions. The wall of shroud were taken to be adiabatic with no slip condition. The main
objective of the analyses was to estimate the mass flow rate incurred in the inlets and outlets of
the system, in order to optimize the shape of the shroud and to evaluate the fan’s capability in
providing the required static pressure.
In order to overcome the computational difficulties and complexities in solving the model
the radiator is consider as a porous zone in the analyses with a porosity of 0.86. Furthermore,
to simplify the fan model, it was computed by providing fan characteristics from the
manufacturer rather than solving for fan data. More fan selection calculations are presented in
the appendix A: 2.
4.7 Advantages and limitations of CFD simulations
Numerical simulations cannot and should not replace real experiments. Both serve the most
exact and economic prediction of flow processes.
Advantages with CFD-
• specific physical boundary conditions or effects can be considered in isolation
• simulations provide at any point measured data unlike experiments
• many flow parameters are gathered, which are not accessible in experiments
• simulations are able to contribute to a greater understanding of the problem than
experiments
• the costs are usually much lower compared to experiments
Disadvantages with CFD-
• errors may occur due to simple flow models or simplified boundary conditions
• possible uncertainties caused by too little computing values per cell and hence
therefore resulting interpolation errors
• computation time may extend for large models
Results and discussions
Three cases were chosen for CFD analysis. The first one involves using a fan of 4000 RPM,
the second one uses fan of 4500 RPM and the last one uses eight electric fans of 24V. Forced
convective heat transfer occurs in all the cases. The reason for choosing such cases was to
correlate the analytical design of heat exchanger (Core). The electric fans were used to find out
their performance in case of a combat vehicle against conventional hydraulic fans.
The following discussion presents the compact heat exchanger’s analytical results and
performance metrics.
The size of the main and auxiliary radiator based on the calculation of analytical design i.e
(1050 x1050 x 140 mm) for main and (1050 x1050 x 34 mm) for auxiliary. They are rated at
(276 ± 10%) kW and (142 ± 10%) kW respectively.
Figure 11- Heat exchanger system resistance curve (analytical)
Figure 12- Heat transfer coefficient vs air flow rate graph (analytical)
Figure (11) clearly shows a direct proportionality among pressure drop and volumetric mass
flow rate. Upon closely observing it can be said that at double the flow rate approximately four
times is the pressure drop in the heat exchanger core. This indicates that flow regime here was
of turbulent nature which is in-sync with the theoretical concepts as well.
In figure (12), doubling the mass flow rate results in a linear increase of 85% in the heat
transfer coefficient in both main and auxiliary radiators which is also in compliance with Piyush
Sabharwall et.al [15] work on effect of mass flow rate on the convective heat transfer
coefficient.
Figure (13) and (14) below shows the mass flow rate of air at radiator end and fan end
respectively. The negative mass flow rate signifies entry of mass into the system and positive
values signify outward flow of mass. It can be seen in the graph that 4000 RPM fan is not able
to achieve the required air mass flow rate of 9.8 kg/s (obtained analytically). The maximum air
flow rate obtained is 7.9 kg/sec (radiator end) and 8.2 kg/sec (fan end). The 4000 RPM fan
selected delivered about approximately 16% less mass flow of air into the radiator, thus it can
be said that this fan is not suitable for cooling this vehicle. This deficiency could be attributed
to the fact that CFD analyses takes up surface roughness, louvered grilles and cowl in
consideration and all the surfaces pose hindrance in air flow path. Thus, a fan with more power
wattage is needed for this cooling pack to work.
Figure 13- Mass flow rate of air at radiator end with 4000 RPM fan
Figure 14-Mass flow rate of air at fan end with 4000 RPM fan
Figure (15) and (16) depicts that the cooling pack will be able to cool the vehicle’s
peripherals effectively at the fan speed of 4500 RPM. This case satisfies the required minimum
criteria of mass flow rate. The observed simulated air mass flow rate at both radiator and fan
end is ~10 kg/s.
Figure 15-Mass flow rate of air at radiator end with 4500 RPM fan
Figure 16-Mass flow rate of air at fan end with 4500 RPM fan
The use of electric fans doesn’t satisfy the requirements of cooling load of vehicle, eight
electric fans were used yet only ~ 2kg/sec of mass flow rate (see figure (17) and (18)) was
observed. Electric fans might be useful in cooling auxiliary power unit or backup system. Major
cooling loads of vehicle needs higher mass flow rate which is not feasible in case of electric
fans, though they consume less power than hydraulic fans. Hence, electric fans cannot be
implemented in the cooling pack of combat vehicle.
Figure 17-Mass flow rate of air at radiator end with electric fans
Figure 18-Mass flow rate of air at fan end with electric fans
Figure 19- Image showing mesh generated for computation
Figure 20- Flow streamlines inside the system
Figure 21- Velocity vector plot
Figure 22-Isometric view of velocity vector plots
with electric fans
Figure 23-Closer view-showing eddy formation inside the cowl
Figure 24-Velocity profile at intake louvered grilles
Conclusions and future scope
The present work focussed on analytically designing the heat exchanger for combat vehicles
and correlating the design using computational modelling, involving porous zones on the
calculations. The following conclusions could be drawn from this work-
1) Combat vehicles operates in harsh and humid atmospheric conditions, hence design
procedures needs to be stringent.
2) It has been observed that preliminary analytical designing is must for heat exchanger
design as it provides required boundary conditions for further computational
modelling.
3) A detailed study on compact heat exchanger has been performed by using the ε –
NTU analytical method.
4) The calculation methodology allowed the evaluation of the cooling system behaviour
and was implemented in a spreadsheet to enable the iteration of the input data.
5) Porous media technique was advantageous as it allowed less complexity in
convergence of solution, also helped in lowering the computational time and
resources.
6) The fan rpm couldn’t be assigned due to unavailability of parasitic losses (pressure
drop) in grilles and other surfaces. Hence, fan of higher rpm was selected to anticipate
these resistances.
7) It has been observed, that Electric fans were not feasible/suitable for cooling the
entire heat load of engine and it’s peripheral, as it require higher amounts of flow rate
to dissipate the heat.
However, for better understanding of correlations between various relevant parameters, the
further scope includes-
1) Porous modelling algorithms can be made more robust by making use of data driven
approach.
2) Surface characteristics of fins can be worked upon to allow them to work on higher
Reynolds number ranges.
References
[1] Rohsenow, W. M., Hartnett, J.P. & Cho, Y.I. (1998) Handbook of Heat Transfer, 3rd Ed. New York:
McGraw Hill.
[2] Çetin, B., Gler, K. G. , & Aksel, M. H. (2017). Computational Modeling of Vehicle Radiators
Using Porous Medium Approach. In S. S. Murshed, & M. M. Lopes (Eds.), Heat Exchangers - Design,
Experiment and Simulation. IntechOpen. https://doi.org/10.5772/66281
[3] Shah, R. K., 1981, Classification of heat exchangers, in Heat Exchangers: Thermal-Hydraulic
Fundamentals and Design, S. Kakac¸, A. E. Bergles, and F. Mayinger, eds., Hemisphere Publishing,
Washington, DC, pp. 9–46
[4] Hayes A.M., Khan J.A., Shaaban A.H., Spearing I.G. The thermal modelling of a matrix heat
exchanger using a porous medium and the thermal non-equilibrium model. Int. J. Therm. Sci. 2008;
47(10):1306–1315
[5] Shah, R. K. (2003). Advances in Automotive Heat Exchanger Technology. SAE Transactions, 112,
631–641. http://www.jstor.org/stable/44745431
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exchanger for vehicle applications. Appl. Therm. Eng. 2005;25:1995-2013.
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[8] Çetin, B., Gler, K. G. , & Aksel, M. H. (2017). Computational Modeling of Vehicle Radiators
Using Porous Medium Approach. In S. S. Murshed, & M. M. Lopes (Eds.), Heat Exchangers-Design,
Experiment and Simulation. IntechOpen. https://doi.org/10.5772/66281
[9] Chiou, J. P. (1975). Engine Cooling System of Military Combat/Tactical Vehicles. SAE
Transactions, 84, 173–191. http://www.jstor.org/stable/44681923
[10] Engineering design handbook- Military vehicle power plant cooling. US Army Material Command,
1975
[11] Kays W M, London A L. Compact Heat Exchangers. 3rd ed. New York: McGraw-Hill; 1984.
[12] Andrew Roberts, Richard Brooks, Philip Shipway, Internal combustion engine cold-start
efficiency: A review of the problem, causes and potential solutions, Energy Conversion and
Management, https://doi.org/10.1016/j.enconman.2014.03.002.
[13] Hussam Jouhara, Navid Khordehgah, Sulaiman Almahmoud, Bertrand Delpech, Amisha Chauhan,
Savvas A. Tassou, Waste heat recovery technologies and applications, Thermal Science and
Engineering Progress, Volume 6, 2018, Pages 268-289, ISSN 2451-9049,
https://doi.org/10.1016/j.tsep.2018.04.017.
[14] Kakaç, S., Liu, H., & Pramuanjaroenkij, A. (2020). Heat Exchangers: Selection, Rating, and
Thermal Design (4th ed.). CRC Press. https://doi.org/10.1201/9780429469862
[15] Piyush Sabharwall, Vivek Utgikar & Fred Gunnerson (2009) Effect of Mass Flow Rate on the
Convective Heat Transfer Coefficient: Analysis for Constant Velocity and Constant Area Case, Nuclear
Technology, 166:2, 197-200, DOI: 10.13182/NT09-A7406
Appendix
A1: Surface parameters of fins
In the present work, louvered fins with surface designation 3/4 - 11.1 was selected for air
side and 1/10-19.74 offset-serrated fins for coolant side were selected. The selection of
following fins was done based on iterating the values until optimum solution is achieved.
All the experimental data for fins and the performance characteristics was taken from Kays
& London. [11]
Table 8- Surface characteristics of selected fins
Surface geometry characteristics
Parameters
Air side
Coolant Side
Type of fin
Louvered fin
Offset serrated fins
Selected Surface
3/4 - 11.1
1/10-19.74
Plate spacing (m)
0.00635
0.00129
Hydraulic radius (m)
0.0007712
0.000305
Fin thickness (m)
0.000152
0.000051
Transfer area/ volume between plates (m
2
/m
3
)
1204.00
3028.00
Fin density (/m)
437.00
777.00
Fin area/total area
0.76
0.51
Figure 25- Colburn and fanning friction factor relation with Reynolds number (a) Louvered
fin (3/4 - 11.1), (b) Offset serrated fin (1/10-19.74) [11]
A2: Fan sizing and selection calculations
Figure 26: Flowchart showing fan selection procedure
Figure 27- Overlay plot of fan curve and system resistance
Based on the overlay plot, AMTEK 475 MP4 axial flow hydraulic fan with outer diameter
as 540 mm and hub diameter as 130 mm was selected for further simulation purpose. This fan
has been used on numerous military combat vehicles viz. Titan, BMP, Rosomak, T-72 and
others.
Electric fan used in the simulation-
The electric fan used in the simulation was SPAL VA15-BP70-/LL-51A brushed axial fan
rated at 24 V.
Figure 28- Characteristics of SPAL electric fan
A3: Excel program tables
Figure 29- Screenshot of Excel tool developed for rating and sizing of auxiliary/main radiator
Calculations and formulas used in the excel tool-
1) Surface characteristics- these should be selected from a known source. Experimental
data is available in Kays & London [11].
2) Calculation of heat transfer and free-flow areas:
(1) Taking heat exchanger core dimensions, calculate frontal areas, A
fr
, on both
sides, and volume, V.
(2) Select material based on inlet temperatures and pressures.
(3) Assume a plate thickness, t, between fluids, which can be verified later.
Heat transfer and free flow areas is given by-
(18)
The total heat transfer area on each side is given by-
(19)
The ratio of free-flow area to frontal area is given by-
(20)
The free-flow areas are then given by-
(21)
3) Fluid properties from standard tables
(1) Assume the heat exchanger effectiveness and estimate an average
temperature for the evaluation of fluid properties.
(2) From the definition of effectiveness Equation 4 for given Cc and Ch,
calculate outlet temperatures.
(3) Now, from standard tables, take viscosity, μ, thermal conductivity, k,
Prandtl number, Pr, specific heat, Cp, at these bulk temperatures.
4) Calculation of Reynolds numbers, using equations as-
(22)
5) Calculation of StPr
2/3
(j factor) and ƒ factor from the basic characteristics of the surface
from Kays & London. [11]
6) Calculation of heat transfer coefficients on both sides using equation 10.
7) Calculation of fin effectiveness on both sides as-
(23)
(24)
8) Calculation of surface effectiveness on both sides as-
(25)
9) Calculation of overall coefficient (U) of heat transfer as-
(26)
10) Calculation of NTU and exchanger effectiveness-
Calculate heat capacity rates and capacity ratio as-
(27)
(28)
11) Calculation of pressure drops on both sides using the complete equation for the pressure
drop as-
(29)
A4: Thermal circuit of the vehicle flowchart
Figure 30- Screenshot of the thermal balance circuit of the excel tool
A5: Gantt chart
Figure 31- Gantt chart describing project timeline
